In high-pressure compressors in particular, sealing against the environment is achieved by means of a shaft seal which is generally formed as a dry gas seal. This seals an inlet pressure against the environment on both axial sides of the compressor. In addition, a compensating piston seal which seals the outlet pressure against the inlet pressure on the pressure side of the compressor is provided to reduce the thrust of the engine and to ensure the inlet pressure on both sides of the shaft in front of the dry gas seal.
Generally, these seals have a hollow stator which embraces the rotor, and the rotor, stator, or both, have recesses on the surfaces. In operation, i.e., when the shaft is rotating, a dynamic resistance is formed between the opposite surfaces of the rotor and stator which opposes a movement of the fluid in axial direction through the sealing gap.
The design of this compensating piston seal is very important for the functionality of the flow machine because the greater pressure difference is generally sealed by this seal and, therefore, the greater dynamic forces occur between the rotor and stator. These dynamic forces influence the stability of the running behavior among other things. When this seal is correctly designed, the rotordynamic stability of turbo compressors can be substantially improved, for example.
Hole pattern (HP) seals in particular are known as a special constructional form of compensating piston seals in which the recesses provided on the inner surface of the stator have the shape of substantially circular holes. In addition, honeycomb (HC) seals are also known in which the recesses provided on the inner surface of the stator are honeycomb-shaped, i.e., have a netlike shaped hexagonal holes. A gap is formed between the inner surface of the stator and the outer surface of the rotor so that there is no contact between the two sealing surfaces.
To ensure the positive effect of the hole pattern design, it is crucially important to be aware of and monitor the geometry of the sealing gap during operation. Formerly, in conventional constructions this was difficult and sometimes impossible. Therefore, compressors with hole pattern seals were often unsuccessful in the past due to rotordynamic instability. The complex of problems will be illustrated in the following example.
FIG. 3 shows a known, in-house compressor 100. An autoclave cover 104, as it is called, is inserted into an outer housing 102, an inner housing 106 being supported at this autoclave cover 104. The housing is closed by a closing cover 108. A shaft 110 is supported by shaft bearings 112 and 112′ in bearing housings 114 and 114′, respectively, which are in turn fastened to the autoclave cover 104 and closing cover 108. The compressor stages with their installed components (not shown in more detail) are located in a work space 116 which is defined by the autoclave cover 104, the inner housing 106, the closing cover 108 and the shaft 110.
Shaft seals 124, 124′ which seal an inlet pressure of the compressor against the ambient pressure are arranged on both sides of the work space. The inlet pressure prevails at the inner compressor side of these two seals so that the shaft seals 124, 124′ are pressed apart by the pressure difference between the inlet pressure and the ambient pressure. Seal spaces at the inner compressor sides of the two shaft seals 124, 124′ communicate with one another via an equalization line (not shown).
In addition, a compensating piston seal 122 is provided on the outlet side (at left in FIG. 3) between the seal space and the actual work space. This compensating piston seal 122 is formed substantially of an end portion 106a of the inner housing 106 and a seal bushing 120 inserted therein and seals the outlet pressure against the inlet pressure.
FIG. 4 shows the area of this compensating piston seal 122 in detail. FIG. 4 is an enlarged view of a detail which is indicated in FIG. 3 by a circle “IV” in dash-dot lines. As is shown in FIG. 4, the work space 116 with its installed parts on the outlet pressure side is defined by the radial and axial inner surfaces of the inner housing 106 and outer surface of the shaft 110. A radially inwardly projecting end portion 106a of the inner housing 106 annularly encloses a sealing portion 110a of the shaft 110 and forms the boundary of the work space 116 in axial direction. A seal element 120 is arranged at the inner surface of the end portion 106a. This seal element 120, which contains the above-described recesses (not shown), reduces the gap between this inner surface of the end portion 106a of the housing and the outer surface of the sealing portion 110a of the shaft to a predetermined extent and defines the geometry of the gap.
The inner housing is formed of two parts, an upper half and a lower half, to allow the rotor to be inserted. The seal element 120 which is formed as a seal bushing is likewise split in radial direction into an upper half and a lower half. These two half-rings are screwed into the corresponding grooves of the inner housing.
However, the seal arrangement described above has some disadvantages. A substantial difficulty with respect to dimensioning and operation is illustrated in FIGS. 5A to 5C. FIGS. 5A to 5C substantially correspond to the section in FIG. 4, but are substantially more schematic. Only portions of the housing 102, autoclave cover 104, inner housing 106, including its end portion 106a which, together with the seal element 120, forms the compensating piston seal 122, and portions of the shaft 110 and work space 116 are shown. A sealing gap between the seal element 120 and the shaft 110 is designated by 140. FIG. 5A shows the geometry, as produced, which represents the design state. FIG. 5B shows the influence of a large, mostly transient, temperature difference between the outer housing and the inner housing on the geometry of the seal arrangement, this temperature difference being based in part on the fact that the inner housing becomes hot substantially faster than the outer housing when the engine is started, and FIG. 5C shows the influence of a large pressure difference along the compensating piston seal 122. FIGS. 5B and 5C show the finished, unloaded geometry from FIG. 5A in dashed lines.
As is shown in FIG. 5A, the sealing gap 140 in hole pattern seals and honeycomb seals in the design state becomes narrower outward, i.e., converges in the assumed flow-out direction or leakage direction. Under the influence of a large temperature difference, the inner housing 106 expands, the end portion 106a expands toward the inside, and the sealing gap 140 becomes narrower (see FIG. 5B). Further, the expansion of the end portion 106a is blocked by a shoulder 104b of the autoclave cover 104 so that the entire end portion 106a rotates around this shoulder 104b. Therefore, the sealing gap 104 not only becomes narrower but is also divergent in addition. Under the influence of a large pressure difference between the outlet pressure and the inlet pressure along the seal, the end portion 106a bulges outward, which also results in the sealing gap 140 becoming more divergent. As a result, the gap geometry is very difficult to control. In extreme cases, this leads to a divergent gap which results in unstable rotordynamics. The change in geometry of the sealing gap 140 can even take on the order of magnitude of the gap height.
The object of the present invention is to improve the compensating piston seal in a flow machine.